1. Field of the Invention
The present invention relates to a method of producing an oil-free screw compressor and, more particularly, to a method of producing an oil-free screw compressor in which the gap between meshing rotors is minimized to improve the compression performance. The invention also relates to the compressor thus manufactured.
2. Description of the Prior Art
An oil-free screw compressor employs a pair of screw rotors which mesh with each other without any lubricating oil existing therebetween. Therefore, a gas under pressure or after compression tends to leak back to the suction side of the compressor through a gap between the meshing rotors, a gap between the rotors and the rotor casing wall and a gap between the discharge side end surfaces of the rotors and the discharge side end cover. Such leakage of the gas to the suction side adversely affects the performance of the screw compressor. In particular, leakage through the gap between the rotors is most critical because this gap forms a seal line between the compression chamber and the suction chamber. Namely, when large pressure differential exists between the suction and discharge chambers across the seal line, the compressed gas flows from the discharge chamber back to the suction chamber at a high rate, thus significantly affecting the performance of the compressor.
Meanwhile, there is a current trend towards reduction in machine size also in the field of oil-free screw compressors. Screw compressors of smaller sizes naturally provide smaller air discharge rate than screw compressors of larger sizes. Nevertheless, the leak of the gas inside the screw compressor is not reduced in proportion to the reduction in the air discharge rate. Attempts therefore have been made to minimize the gap between the rotors so as to reduce the reverse flow of the high-pressure gas from the discharge side to the suction side of the compressor.
Single-stage oil-free screw compressors are broadly used for compressing gases from atmospheric pressure to 7 ata or so. In such compressors, the compressed gas exhibits a high temperature well exceeding 300.degree. C., so that the lobe portions of the rotors are thermally expanded and deformed by the heat of the compressed gas. It is therefore necessary that the profiles of the lobe portions of the rotors be designed in full consideration of the thermal expansions of the rotors.
To cope with such a demand, a method of producing a compressor is disclosed in, for example, Japanese Patent Examined Publication No. 61-47992. In this method, as shown in FIG. 5 which corresponds to FIG. 12 of the above-mentioned Japanese patent publication, a profile of a pair of rotors which mesh with each other at normal temperature are used as a basic rotor profile. Then, a first rotor profile is determined for one of the rotors which is assumed here to be a male rotor, taking into account the thermal expansion which is predicted to occur when this rotor is heated to a predetermined maximum credible temperature. Then, a second rotor profile is determined for the male rotor taking into account the back lash between the pair of rotors and an ideal gap size which would not cause the meshing rotors to contact with each other during operation. Then, a third rotor profile is determined for the rotor meshing with the male rotor, i.e., a female rotor. The third rotor profile is a profile which is generated based on the second rotor profile and which is to be assumed by the female rotor when the latter is deformed by thermal expansion. Then, a fourth rotor profile is determined for the female rotor. This fourth rotor profile is a profile to be exhibited by the female rotor when the female rotor, which assumes the third rotor profile in expanded state, is contracted by being cooled down to normal temperature. Finally, the rotors are fabricated at normal temperature based upon the fourth rotor profile for the female rotor and the basic rotor profile for the male rotor.
In oil-free screw compressors, the rotors rotate at very high speeds, e.g., 60 to 100 m/s in terms of the peripheral velocity. Therefore, any inadequate design value based upon prediction may lead to a serious accident such as breakdown of the rotors due to interference between the rotors. Damaging of the rotors may occur also when the compressor operates at a temperature which falls out of the range of given specifications. For instance, if the suction pressure is decreased down below a predetermined set pressure while the discharge pressure is elevated to exceed a set pressure, the compression ratio exceeds the set value so that the compressed gas such as air exhibits a temperature higher than the design temperature. In such a case, the rotors are expanded beyond limits to interfere with each other and thus are damaged. Besides the predicted temperature rise of the rotor as described, the following factors affect the size of the gap between the rotors:
(1) Each rotor exhibits a temperature distribution also in each cross-sectional plane perpendicular to the rotor axis. It is very difficult to exactly predict the amount of the thermal expansion of the rotor which has three-dimensional twisted form by accurately grasping temperature distributions in all such cross-sectional planes.
(2) The rotor temperature continuously varies also in the axial direction. Namely, the rotor has temperature distribution also in the axial direction. Practically, there is a temperature differential of 100.degree. to 200.degree. C. between the discharge side and the suction side. In view of such temperature differential, the rotor profile is designed with a certain degree of taper imparted to the profile surface. Practically, this tapered design is achieved merely by axially shifting the profile determined for the discharge end so as to simulate the rotor profile at each axial position with the above-mentioned discharge end profile. Ideally, however, the rotor profile should be determined by assuming numerous axial points over the entire rotor length, grasping the temperature distribution in the cross-section at each of such points, determining an ideal rotor cross-sectional shape for each of the cross-sectional planes, and smoothly connecting such ideal cross-sectional shapes of the numerous axial points. Such a complicated rotor configuration, however, can never be achieved with a known machine such as a hobbing machine or a gear teeth grinder nor by a machine such as a single index cutter having a cutting edge which is shaped to conform with a fixed rotor profile.
(3) Machining error exists not only in the production of rotors but also in the fabrication of the casing. Furthermore, the casing also has different temperature distributions at the suction and discharge sides thereof. Consequently, the distance between the rotor axes is changed during operation to affect the size of the gap between the rotors. It is extremely difficult to estimate the amount of such a change in the rotor gap size.
(4) A gap also exists between each end of the rotor and a bearing supporting the rotor end. The size of this gap also varies during operation, thus forming one of the factors which cause change in the distance between the rotor axes.
(5) In operation of the screw compressor, the gas compressed between rotating rotors causes deflection of the rotors. The deflection is distributed three-dimensionally. Designing the rotor configurations taking into account such rotor deflection is a highly complicated work and cannot easily be carried out.
(6) There are also other factors which affect the size of the gap between the rotors, such as difference in dimensions between individual screw compressors incurred in the course of production, variation in the operating conditions, and so forth. Thus, it is extremely difficult to precisely control extremely small size of the gap between the rotors of screw compressors.